JPS5865995A - Compressor - Google Patents

Abstract

PURPOSE:To secure pertinent capacity control characteristics in conformity with the running properties of engine vehicles, by setting up two flow passages different in the position of each nozzle opening leading to a vane chamber, while making effective areas of these flow passages selectable at will, in a rotary compressor. CONSTITUTION:As made known when having an eye to a vane chamber 26a shown in illustration, the influx route of refrigerant into the vane chamber 26a varies, as in illustrations (a) thru (e), with a vane traveling angle theta, during suction stroke. In case of small cars being subject to high duty frequency at low- speed running and especially required to meet low-power consumption, it will do merely in the following way that a spacer 52 which is large in the opening area of a nozzle 54 and relatively small in the opening area of a nozzle 53 is selected. Also in case of large and heavy-duty cars equipped with a large capacity compressor and required to meet energy-saving effects, it is only enough to select the spacer 52 being smaller in the opening area of the nozzle 54 than that of the nozzle 53.

Description

[Detailed Description of the Invention] The present invention provides an air conditioner 2 using a rotary compressor.
-- Concerns the control of refrigeration capacity in the system.

Before explaining the present invention, a sliding vane type rotary compressor for a car cooler will be explained.

As shown in Fig. 1, a general sliding vane type compressor consists of a cylinder 1 having a cylindrical space inside, and a blade chamber 2 fixed to both sides of the cylinder and serving as an internal space of the cylinder 1.
a side plate (not shown in FIG. 1) that seals the cylinder 1 on its side, and a rotor 3 eccentrically arranged within the cylinder 1.
The vane 5 is slidably engaged with a groove 4 provided in the rotor 3 of and. 6 is a suction hole formed in the side plate, and 7 is a discharge hole formed in the cylinder 1. vane 5
As the rotor 3 rotates, it flies outward due to centrifugal force, and its tip surface slides on the inner wall surface of the cylinder 1, thereby preventing gas from leaking from the compressor.

Such a sliding vane type rotary compressor has a complicated structure and is a reciprocating compressor with a large number of parts, and has recently been applied to compressors for car coolers. However, this rotary set had the following problems compared to the reciprocating type.

In other words, in the case of a car cooler, the driving force of the engine is
It is transmitted to the clutch pulley via the belt and drives the rotating shaft of the compressor. Therefore, when a sliding vane compressor is used, its refrigerating capacity increases almost linearly in proportion to the number of revolutions of the car engine.

On the other hand, when using a conventionally used reciprocating type compressor, the followability of the suction valve becomes poor at high speed rotation, and the compressed gas cannot be sufficiently sucked into the cylinder, resulting in a decrease in refrigeration capacity. At high speeds, it becomes saturated and shuffles. With the shunting and reciprocating type, the effect of suppressing the refrigerating capacity automatically works when running at high speed, but with the rotary set, this effect does not work, and the efficiency decreases due to increased compression work, or overcooling (cooling) occurs. (too much). As a method for solving the above-mentioned problems of the rotary compressor, a control valve that changes the opening area of the flow path is configured in the flow path leading to the suction hole 6 of the rotary compressor,
Conventionally, a method has been proposed in which the opening area is narrowed during high-speed rotation and the suction loss is utilized to increase the capacity. However, in this case, the above-mentioned control pulp must be added separately, resulting in a problem that the structure becomes different and the cost increases.

As another method for solving the problem of excessive capacity of a rotary compressor at high speeds, a structure has been proposed that uses a fluid clutch, a planetary gear, etc. to prevent the rotation speed from increasing above a certain level.

However, for example, the former has a large energy loss due to frictional heat generated by the relative moving surfaces, and the latter has a large size and shape due to the addition of a planetary gear mechanism with many parts. Nowadays, there is a demand for compactness, but it is difficult to put it into practical use.

In order to solve the above-mentioned problems associated with the a-tally conversion of the refrigeration cycle for car coolers, the present inventors have determined that the impeller chamber 6 when a rotary compressor is used is the result of a detailed study of pressure transient phenomena. Even in the case of a rotary compressor, its suction hole area and discharge volume.

By appropriately selecting and combining parameters such as the number of blades, we have discovered that, like the conventional reciprocating system, the self-suppressing effect of the refrigerating capacity during high-speed rotation works effectively. The application is currently being filed under No.

Furthermore, based on the results of consideration regarding the overall characteristics of the compressor, which takes into account not only the volumetric efficiency but also the power consumption, the inventors created a multi-stage configuration for changing the effective suction area, and set the effective area in the first half and the second half appropriately. It has been discovered that by doing so, it is possible to reduce the driving torque at low speed rotations, and to obtain a sufficient capacity control effect even at high speed rotations. .

In the present invention, at least two flow passages are formed with different opening positions in the cylinder in the circumferential direction, and the effective areas of these two flow passages can be arbitrarily selected.

For example, by attaching a spacer to the flow path, it is possible to easily obtain a compressor with appropriate performance characteristics or torque characteristics according to the characteristics of the engine vehicle, which has already been proposed. Special request made in 1982
This invention is an improvement on the invention according to No. 62875 from the aspect of compressor performance.

Hereinafter, as an example, a case will be described in which the present invention is applied to a two-vane type sliding vane compressor.

Hereinafter, the present invention will be explained in the following order.

Explanation of the present invention using a two-vein compressor as an example m Explanation of the principle of capacity control that is the basis of the present invention [TIl] Supplementary explanation Method for measuring the effective suction area First, the case of [I] will be explained below. do.

FIG. 2 is a front sectional view of a compressor showing one embodiment of the present invention, and FIG. 3 is a side view.

11 is a cylinder, 12 is a low pressure side blade chamber, 13 is a high pressure side blade chamber, 14 is a vane, 15 is a sliding groove of the vane, 16 is a rotor, 19 is a discharge hole, 200 is a head cover, and 202 is a suction pipe joint. .

50 is a suction hole A formed in the cylinder 11, 61 is a suction hole B, 52 is a spacer, 53 is a spacer 52 and a spacer 5.
Bolts 66 are formed inside the helad cover 200, 57 are bolts for attaching the helad cover 200 to the cylinder 11, 5
Reference numeral 8 denotes a cast hole formed in the inner surface of the cylinder 11 during mold processing.

In Fig. 3, 2o is a front panel which is a side plate;
1 is a rear panel, 22 is a rotating shaft, 23 is a rear case, 24 is a disk of the clutch fixed to the rotating shaft 22, and 25 is a pulley.

With the compressor of the present invention, even after removing only the spatula cover 200, the suction hole area of 1, t (nozzle A63. A spacer 52 having an area of Be5a) can be attached and detached.

In the present invention, by selecting the spacer 52, even if compressors having the same configuration except for the spacer 52 are used, a compressor having capacity control characteristics that matches various engine and vehicle characteristics can be easily and arbitrarily created. The feature is that it can be configured.

Now, the compressor in the embodiment is constructed under the following conditions.

Table 1 Angle in Table 1 = 01. θ2. θ6 is defined as follows.

That is, in FIG. 4, 26-a is the blade chamber A.

97--2 26-b is the blade chamber B, 27 is the top part of the cylinder 11,
28-a is the bee 7A, 28-b is the vane B, and 29 is the end of the suction hole Aao.

The position where the tip of the vane passes through the top portion 27 of the cylinder with the center of rotation of the rotor 16 as the center is θ-0, and the angle at any position of the tip of the vane is θ with the θ=◯ as the origin. Focusing on the blade chamber A26-a, during the suction stroke, depending on the vane running angle: θ, the blade chamber A26-a
The inflow path of the refrigerant to a changes as shown in FIG. 4 A to E. In addition, suction hole A50. FIG. 5 shows an example of the effective area of the inflow path of the blade chamber A26-a as seen from the common inlet of the suction hole B51: a.

In Figures I to E of Fig. 4, Figure I shows the state immediately after the vane A28-a passes the top portion 27, and immediately after the suction stroke has started.

The second time, the vane A28-a connects the suction hole B61 and the suction hole At.
5o, and at this time the blade chamber A
Refrigerant is supplied to 26-a only from the suction hole B51.

561 The effective area at this time is decided by the nozzle B54,
In the case of using the suction effective area pattern A (FIG. 5), which is one of the embodiments, a=a2-0.185 cdl.

FIG. 4C shows that the vane A25a passes through the suction hole A6o,
A case is shown in which the vane B28-b is in a state in front of the suction hole B51.

At this time, refrigerant is supplied to the blade chamber A26-a from both the suction hole A50 and the suction hole B51.

Pattern A' (y: 1n) In the example, nozzle A33
(7) Effective area: a = a 1 = 0,416 cl, therefore, if refrigerant is supplied from both nozzle A53 and nozzle B51, the total effective area: a ”” a
1+a 2-〇, 6crl.

FIG. 42 shows a state in which the vane B28-b is located between the suction hole B51 and the suction hole A6o.

At this time, the blade chamber A26-a has a suction hole Aeso
The refrigerant is supplied from only the suction hole B61, and the refrigerant is supplied from the suction hole B61.
It is blocked by vanes B28-bK.

The effective area in the above section is 11 only for the nozzle At53.
- determined by a=a, =0.4150J.

Figure E shows the state immediately after the suction stroke of the vane chamber A2e-a is completed, and the tip of the vane B28-b is located at the suction hole end 29. At this point, the suction hole A30i was formed so that the volume of the blade chamber partitioned by the vane A28a and the vane B25b was maximized.

Now, in the compressor according to the present invention, for example, the spacer 62
f: By providing it in the suction flow path, a compressor having arbitrary characteristics matching the characteristics of the engine and vehicle can be easily obtained using the same compressor.

The characteristic of the compressor that can be selected is that not only the capacity control effect but also the torque (power consumption) characteristics can be considered.

For example, in the case of a small car that is frequently used in low-speed operation and requires particularly low power consumption, the spacer 62 with a large opening area of the nozzle B54 and a relatively small opening area of the nozzle A53 within an appropriate range may be used. Just choose.

Alternatively, if the vehicle is a large vehicle equipped with a large-capacity compressor and a large energy-saving effect is required at high speed rotation, the nozzle B5 is suitable for obtaining sufficient capacity control effect at high speed rotation.
4, the spacer 52 formed with a smaller opening area than the nozzle A53 may be selected.

As will be described later, the closer the effective suction area is to a constant value during the suction stroke, the greater the effect of suppressing the refrigerating capacity at high speeds. Nozzle A
The reason why the capacity control characteristics and torque characteristics of the compressor change depending on the selection of the effective area of the nozzle B 63 and the nozzle B 54 will be explained in detail in [Ir'].

FIG. 6 shows another embodiment of the present invention, in which 150 is a rotor, 151 is a vane, 152 is a head cover, and 15
3 is a cylinder, and 164 is a suction hole A.

165 is suction hole B, 156 is spacer A, 157 is nozzle A formed in spacer A, and 158 is spacer A.
169 is a spacer B, 1 s.
o is a nozzle B, and 161 is a threaded portion.

In the above embodiment, unlike the embodiment shown in FIG.
--Separate spacers are provided for forming nozzle A156 and nozzle B160'z, and both spacers 1
56 and 159 are provided in the cylinder 163 with threaded portions.

[Old] Capacity control using a multi-stage suction configuration, which is the basis of the present invention, will be explained below.

The transient characteristics of the blade chamber pressure can be described by the following energy equation.

In the above equation, G: weight flow rate of refrigerant, ■ a = blade chamber volume, A e heat equivalent of work, CP: specific heat at constant pressure, TA: supply side refrigerant temperature, N: specific heat ratio, R: gas constant, C ■: Specific heat at constant volume, Pa: pressure in the blade chamber, Q: amount of heat, γa = specific weight of the refrigerant in the blade chamber, Ta: temperature of the refrigerant in the blade chamber. In the following equations 2 to 4, a: effective area of the suction hole, q: gravitational acceleration, γA: specific weight of the refrigerant on the supply side, and PS: pressure of the refrigerant on the supply side.

In Equation 1, the first term on the left side is the thermal energy of the refrigerant that passes through the suction hole and is brought into the blade chamber per unit time, and the second term is the work done by the refrigerant pressure on the outside per unit time. The third term represents the thermal energy that flows in from the outside through the outer wall per unit time, and the right side represents the increase in the internal energy of the system per unit time. Assuming that the refrigerant follows the ideal gas law, and since the suction stroke of the shredded compressor is rapid, it changes adiabatically. From Q 1, = P a /RT a, iT = O, the following equation is obtained. .

The nozzle theory can be applied to the weight flow rate of the refrigerant passing through the suction hole, so by solving equations 3 and 4 simultaneously,
The transient characteristics of the blade chamber pressure Pa can be obtained.

When 0〈θ〈π, Va(θ)=■(θ)π〈θ〈θS
When, Va (θ) = ■ (θ) - V (π - θ) Formula 5 Above: ΔV (θ) is a correction term due to the fact that the vane is eccentrically arranged with respect to the rotor center. is usually on the order of 1-2%.

Now, the pattern of the effective suction area during the suction stroke is modeled and analyzed as shown in Figure 7 and Table 2.

The actual effective area shows complex changes as shown in Figure 5, and differs from the monotonous pattern shown in Figure 7, but the purpose of this analysis is to determine the size of the effective area in the first and second half of the suction stroke, and the compression This is done to understand the correlation between each characteristic of the machine.

Table 2 Figure 8 shows the rotation speed as a parameter under the initial conditions of t = O, P = P s, using the conditions of Equations 3 to 5 and Table 19, Table 2, C0, and Table 3. The transient characteristics of the blade chamber pressure in the case of effective suction area C in Figure 7 are obtained. Also,
The refrigerant in the car cooler refrigeration cycle usually uses R12, so -1,13°R=668Ay・cnVloK
#, γA=16.8X10-6#/en.

The analysis was conducted with the angle TA” set at 283°.

In Figure 8, at low speed rotation Cω-100 Or pm)
Then, around θ-2700 when the suction stroke ends, the blade chamber pressure Pa has already reached the supply pressure: Pa-3, 1sA1r/
cJabs177.- has been reached, and there is no loss of blade chamber pressure at the end of the suction stroke. When the rotation speed increases, the refrigerant supply cannot keep up with the change in the volume of the blade chamber, and at the end of the suction stroke (θ-2
7o0) increases. For example, ω
= 5000 rpm, the pressure loss with respect to the supply pressure: PS is ΔP = 1.30@/lri, which results in a decrease in the total weight of suction refrigerant, resulting in a significant decrease in refrigerating capacity.

When the effective suction area is as shown in Figure 7 and when the effective suction area is as shown in Figure 7.
The case of Figure A is shown in Figures 9 and 10, respectively.

Now, the blade chamber pressure at the end of the suction stroke is Pa=Pa
, s, then the pressure drop rate: ηp can be considered to be approximately equal to the drop rate of the total weight of refrigerant filled into the blade chamber at the end of the suction stroke.

Figure 11 shows the case where the effective suction area is different (7th
It is a graph which shows the characteristic of the said pressure drop rate: (eta)p with respect to the rotation speed of (a)-(a) of a figure.

That is, (1) At low speed: ω = 200 Or p m, the pressure drop rates of the compressors having effective suction areas I to I are almost the same.

■ At high speed: ω = 500 Or pm, the compressor (a) whose effective suction area is constant during the suction stroke has the largest pressure drop rate.

■ In the example of Table 1 where the effective suction area is C,
In a compressor with almost the same characteristics as A above, and with a suction effective area, ηp, which is the effect of capacity control, was quite small.

197, --- Figure 12 shows the effective area of the discharge port as a-0, 21c++
1 shows the characteristics of the driving torque with respect to the rotational speed when the effective suction areas are different (A to A in FIG. 7).

Now, when the capacity side (goods) is applied, the internal division of the drive torque of the compressor is as follows.

■Loss in the suction stroke■Compression power in the compression stroke■Loss due to overcompression pressure Regarding the above ■~■, see the model diagrams in Figure 13 and
This will be explained using Figure 4.

In FIG. 13, the curve drawn by abed: N1 indicates the standard redrop suction compression stroke of the compressor.

Curve drawn by ab'efgd: N2 is the case when capacity control is applied, and the effective suction area is constant during the suction stroke, for example, Pv when it has the effective area shown in Figure 7 A.
It is a line diagram.

When the capacity side (goods) is applied, the blade chamber pressure Pa at the start point of the compression stroke decreases as the rotation speed increases.

When capacity control is not performed, the compression stroke start point; Va=
The blade chamber pressure Pa at 47cc (or the end point of the suction stroke) indicates overcompression pressure regardless of the rotation speed.

The curve N3 in FIG. 14 is a PV diagram corresponding to the opening to the end of FIG. 7 in which the effective suction area has a two-stage configuration.

Area: 81 is the loss of power during the suction stroke. Area: 82 is the reduction in compression power due to the capacity side leg effect. Area: 83 is the loss of overcompression power.

When the effective suction area is constant during the suction stroke (FIG. 7A), the blade chamber pressure pressure Pa must start to drop while the blade chamber volume: va is small, and the loss dynamic pressure crab 81 is large. On the other hand, when the effective suction area is large in the first half of the suction stroke and becomes small in the second half (for example, Fig. 7C), the drop in the blade chamber pressure Pa is small in the first half, so the suction loss as a whole is:
S is smaller than the former.

Figures 15 and 16 show the breakdown of the suction loss and overcompression loss to the 21 Beni I~ for each rotational speed. It can be seen that the smaller the change in the effective suction area during the suction stroke, the greater the suction loss, and conversely the greater the overcompression loss.

From the above analysis results, it has been found that by appropriately setting the size of the effective area in the first and second half of the suction stroke, the most favorable characteristics can be selected for the compressor and the vehicle equipped with this compressor.

This compressor focuses on this point, and by forming at least two suction flow passages with spacers, the effective suction area can be arbitrarily selected without disassembling other parts of the compressor. It was made in a similar manner.

[110 Now, the effective area for inhalation in the present invention is as follows.

If there is a point in the fluid path from the evaporator outlet to the compressor blade chamber where its cross-sectional area is the smallest,
From the value obtained by multiplying the cross-sectional area by the contraction coefficient: C=0.7 to 0.9, the approximate value of the effective suction area: a can be determined. however,
Strictly speaking, the effective suction area: a is defined as the value obtained from the following experiment according to the method used in JIS B832022, etc.

Figure 17 shows an example of the experimental method.
is a compressor, IQl is a high-pressure air supply pipe connected from the evaporator to the intake hole of the compressor when installed in a vehicle, Pine-102 is a pipe for supplying high-pressure air, 103 is a housing for connecting both pipes 101 and 102, and 104 is a heat pipe. 105 is a flow meter, 106 is a pressure gauge, 107 is a pressure regulating valve, 108
is a high pressure air source.

The part surrounded by the dot-dash line N in FIG. 17 corresponds to the compressor to which the present invention is applied.

However, in the above experimental apparatus, if there is a constriction part inside the evaporator that cannot be ignored as fluid resistance, it is necessary to add a constriction corresponding to the constriction part to the pipe 1o1.

Now, for example, when measuring the effective suction area a of a compressor with a structure as shown in Fig. 3, remove the clutch disc and pulleys 24 and 25, and
The cylinder 11 should be removed 23,.

You can perform the experiment in this state.

The pressure of the high pressure air source is P1 and the atmospheric pressure is P2.
=1. o3H7crrl a b s, specific heat ratio of air: 1-1.4, specific weight: γ, gravitational acceleration” q=98
If the weight flow rate obtained under the two conditions is G1, then the effective suction area: a can be obtained as shown below.

24 formula However, 0.528〈P2/P1〈○. Set the high pressure: Pl so that it falls within the range of 9.

As described above, according to the present invention, which has at least two intake flow passages and is configured such that the effective area of the flow passage can be arbitrarily selected, it is possible to obtain appropriate torque characteristics according to the characteristics of the engine and vehicle. ) can be easily obtained, and its effects are extremely large.

[Brief explanation of the drawing]

Fig. 1 is a front sectional view of a general sliding vane type rotary compressor, Fig. 2 is a front sectional view of a rotary compressor that is an embodiment of the present invention, and Fig. 3 is a side sectional view of the same compressor. Figure 4 shows the positional relationship between each vane and each suction hole during the suction stroke, Figure 5 shows two patterns of the effective suction area of this compressor, and Figure 6 shows the A sectional view of another embodiment of the present invention, FIG. 7 is a diagram showing the modeled effective suction area, FIGS. 8 to 10 are diagrams showing pressure changes in the suction stroke, and FIG. 11 is a pressure drop rate. 12 is a graph showing torque characteristics, 13 is a graph showing torque characteristics, 14 is a PV diagram, 15 is a graph showing suction loss,
Figure 16 shows overcompression loss for 11-, 1... cylinder,
12・Excellent...Blade chamber, 14...Vane 1
6...Rotor, 19, ψ...Discharge hole, 21
．． 22...Side plate, 50.51...Suction hole, 52...Opening area variable part. Name of agent: Patent attorney Toshio Nakao and 1 other person @1
Figure 2 Figure 4 (8 @ 4 Figure 6 Figure 7 - Shita line shallow 03 (d Cow Figure 11 Crazy 0 Middle heaven number /At (Jprr+) Figure 12 Circulation soft ^ (Oprn) Fig. 14. Wing chamber volume direct cc Fig. 15

Claims (1)

[Scope of Claims] A rotor on which vanes are slidably provided, a cylinder that accommodates the rotor and the vanes, and blades fixed to both sides of the cylinder and formed by the vanes, the rotor, and the cylinder. It is composed of a side plate that seals the space of the chamber on its side, and a suction hole and a discharge hole that are flow passages that communicate the blade chamber with the outside, and the blade chamber pressure in the suction stroke is lower than the refrigerant supply source pressure. In a compressor that suppresses refrigeration capacity during high-speed 1 operation by utilizing suction loss caused by -1 A compressor comprising a variable opening area section provided in the suction flow path for adjusting the refrigerating capacity characteristics.